U.S. patent application number 12/250514 was filed with the patent office on 2009-04-30 for method and apparatus to control clutch fill pressure in an electro-mechanical transmission.
This patent application is currently assigned to GM GLOBAL TECHNOLOGY OPERATIONS, INC.. Invention is credited to Ali K. Naqvi, Jy-Jen F. Sah.
Application Number | 20090111643 12/250514 |
Document ID | / |
Family ID | 40289259 |
Filed Date | 2009-04-30 |
United States Patent
Application |
20090111643 |
Kind Code |
A1 |
Sah; Jy-Jen F. ; et
al. |
April 30, 2009 |
METHOD AND APPARATUS TO CONTROL CLUTCH FILL PRESSURE IN AN
ELECTRO-MECHANICAL TRANSMISSION
Abstract
A method to control a powertrain including a transmission, an
engine, and an electric machine includes applying through a series
of clutch fill events a series of incrementally changing command
pressures in a pressure control solenoid controllably connected to
a clutch within the transmission, monitoring a pressure switch
fluidly connected to the pressure control solenoid and configured
to indicate when the pressure switch is in a full feed state,
determining changes in cycle times of the pressure switch
corresponding to sequential applications of the series of
incrementally changing command pressures, selecting a preferred
command pressure to achieve a transient state in the clutch based
upon the changes in pressure switch cycle times, and controlling
the clutch based upon the preferred command pressure.
Inventors: |
Sah; Jy-Jen F.; (West
Bloomfield, MI) ; Naqvi; Ali K.; (White Lake,
MI) |
Correspondence
Address: |
CICHOSZ & CICHOSZ, PLLC
129 E. COMMERCE
MILFORD
MI
48381
US
|
Assignee: |
GM GLOBAL TECHNOLOGY OPERATIONS,
INC.
Detroit
MI
Daimler AG
Stuttgart
MI
Chrysler LLC
Auburn Hills
Bayerishe Motoren Werke Aktiengesellschaft
Munchen
|
Family ID: |
40289259 |
Appl. No.: |
12/250514 |
Filed: |
October 13, 2008 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
60982874 |
Oct 26, 2007 |
|
|
|
Current U.S.
Class: |
477/5 ; 180/65.7;
701/54 |
Current CPC
Class: |
F16D 2500/50293
20130101; F16D 2500/70605 20130101; Y10T 477/26 20150115; F16D
2500/3166 20130101; F16D 48/066 20130101; F16D 2500/3024 20130101;
F16D 2500/50251 20130101; Y02T 10/6239 20130101; F16D 2048/0221
20130101; Y02T 10/62 20130101; F16H 61/0206 20130101; F16D 48/02
20130101; Y10T 477/69395 20150115; F16H 2061/064 20130101; B60K
6/445 20130101 |
Class at
Publication: |
477/5 ; 701/54;
180/65.7 |
International
Class: |
B60W 10/08 20060101
B60W010/08; G06F 17/00 20060101 G06F017/00 |
Claims
1. Method for controlling a powertrain comprising an
electro-mechanical transmission mechanically-operatively coupled to
an internal combustion engine and an electric machine adapted to
selectively transmit mechanical power to an output member via
selective application of a plurality of hydraulically-applied
torque transfer clutches, said method comprising: applying through
a series of clutch fill events a series of incrementally changing
command pressures in a pressure control solenoid controllably
connected to one of said clutches; monitoring a pressure switch
fluidly connected to said pressure control solenoid and configured
to indicate when said pressure switch is in a full feed state;
determining changes in cycle times of said pressure switch
corresponding to sequential applications of said series of
incrementally changing command pressures; selecting a preferred
command pressure to achieve a transient state in said clutch based
upon said changes in pressure switch cycle times; and controlling
said clutch based upon said preferred command pressure.
2. The method of claim 1, wherein applying through said series of
clutch fill events said series of incrementally changing command
pressures comprises: applying a series of incrementally decreasing
command pressures.
3. The method of claim 1, wherein selecting said preferred command
pressure based upon said changes in pressure switch cycle times
comprises: selecting a lowest of said command pressures conforming
with a subset of shortest of said changes in cycle times of said
pressure switch.
4. The method of claim 1, wherein selecting said preferred command
pressure based upon said changes in pressure switch cycle times
comprises: identifying a first of said command pressures to create
an overlap state to correspond to a change in cycle times of said
pressure switch not conforming with a subset of shortest of said
changes in cycle times of said pressure switch; and selecting as
said preferred command pressure a value equaling the sum of said
identified first of said command pressures and a command pressure
adjustment.
5. The method of claim 4, wherein said command pressure adjustment
equals one and a half times said incremental change utilized in
said series of incrementally changing command pressures.
6. The method of claim 1, further comprising: monitoring said
controlling said clutch based upon said preferred command pressure;
and if said monitoring said controlling ceases to validate said
preferred command pressure, selecting a new preferred command
pressure.
7. The method of claim 1, wherein selecting said preferred command
pressure based upon said changes in said pressure switch cycle
times comprises selecting said preferred command pressure based
upon achieving a touching state in said clutch.
8. Method for controlling a powertrain comprising an
electro-mechanical transmission mechanically-operatively coupled to
an internal combustion engine and an electric machine adapted to
selectively transmit mechanical power to an output member via
selective application of a plurality of hydraulically-applied
torque transfer clutches, said method comprising: selecting a
preferred command pressure in a pressure control solenoid fluidly
connected to one of said clutches, said selecting comprising
iteratively filling said clutch with a calibration command
pressure, wherein said calibration command pressure decreases with
each iteration, monitoring pressure switch cycle times for a
pressure switch fluidly connected to said pressure control solenoid
and configured to generate an indication when said pressure control
solenoid is in a full feed state, and selecting from said
calibration command pressures said preferred command pressure based
upon identifying a transition between differences in said
calibration command pressures; and utilizing said preferred command
pressure in subsequent operation of said clutch.
9. The method of claim 8, further comprising: comparing pressure
switch cycles times during said subsequent operation of said clutch
to expected cycle times; and reselecting said preferred command
pressure based upon said comparing.
10. An apparatus for controlling a powertrain comprising an
electro-mechanical transmission mechanically-operatively coupled to
an internal combustion engine and an electric machine adapted to
selectively transmit mechanical power to an output member via
selective application of a plurality of hydraulically-applied
torque transfer clutches, said method comprising: a pressure
control solenoid fluidly connected to one of said clutches; a
pressure switch fluidly connected to said pressure control solenoid
monitoring when said pressure control solenoid is in a full feed
state; a pressure control solenoid control module, including
programming to apply through a series of clutch fill events a
series of incrementally changing command pressures in said pressure
control solenoid to create an overlap state in said pressure
control solenoid, monitor said pressure switch, determine changes
in cycle times of said pressure switch corresponding to sequential
applications of said series of incrementally changing command
pressures, select a preferred command pressure based upon said
changes in pressure switch cycle times, and control said clutch
based upon said preferred command pressure.
10. The apparatus of claim 9, wherein said series of incrementally
changing command pressures comprises a series of incrementally
decreasing command pressures.
11. The apparatus of claim 10, wherein said programming to select
said preferred command pressure based upon said changes in pressure
switch cycle times comprises: selection of a lowest of said command
pressures conforming with a subset of shortest of said changes in
cycle times of said pressure switch.
12. The apparatus of claim 9, wherein said programming to select
said preferred command pressure based upon said changes in pressure
switch cycle times is based upon determining achieving a touching
state in said clutch.
13. The apparatus of claim 9, wherein said programming further
comprises: validation of said preferred command pressure through
subsequent operation of said clutch; and reselection of said
preferred command pressure based upon a failure in said validation.
Description
CROSS REFERENCE TO RELATED APPLICATIONS
[0001] This application claims the benefit of U.S. Provisional
Application No. 60/982,874 filed on Oct. 26, 2007 which is hereby
incorporated herein by reference.
TECHNICAL FIELD
[0002] This disclosure pertains to control systems for
electro-mechanical transmissions.
BACKGROUND
[0003] The statements in this section merely provide background
information related to the present disclosure and may not
constitute prior art.
[0004] Known powertrain architectures include torque-generative
devices, including internal combustion engines and electric
machines, which transmit torque through a transmission device to an
output member. One exemplary powertrain includes a two-mode,
compound-split, electro-mechanical transmission which utilizes an
input member for receiving motive torque from a prime mover power
source, preferably an internal combustion engine, and an output
member. The output member can be operatively connected to a
driveline for a motor vehicle for transmitting tractive torque
thereto. Electric machines, operative as motors or generators,
generate a torque input to the transmission, independently of a
torque input from the internal combustion engine. The electric
machines may transform vehicle kinetic energy, transmitted through
the vehicle driveline, to electrical energy that is storable in an
electrical energy storage device. A control system monitors various
inputs from the vehicle and the operator and provides operational
control of the powertrain, including controlling transmission
operating state and gear shifting, controlling the
torque-generative devices, and regulating the electrical power
interchange among the electrical energy storage device and the
electric machines to manage outputs of the transmission, including
torque and rotational speed. A hydraulic control system is known to
provide pressurized hydraulic oil for a number of functions
throughout the powertrain.
[0005] Operation of the above devices within a hybrid powertrain
vehicle require management of numerous torque bearing shafts or
devices representing connections to the above mentioned engine,
electrical machines, and driveline. Input torque from the engine
and input torque from the electric machine or electric machines can
be applied individually or cooperatively to provide output torque.
Various control schemes and operational connections between the
various aforementioned components of the hybrid drive system are
known, and the control system must be able to engage to and
disengage the various components from the transmission in order to
perform the functions of the hybrid powertrain system. Engagement
and disengagement are known to be accomplished within the
transmission by employing selectively operable clutches.
[0006] Clutches are devices well known in the art for engaging and
disengaging shafts including the management of rotational velocity
and torque differences between the shafts. Clutches are known in a
variety of designs and control methods. One known type of clutch is
a mechanical clutch operating by separating or joining two
connective surfaces, for instance, clutch plates, operating, when
joined, to apply frictional torque to each other. One control
method for operating such a mechanical clutch includes applying the
hydraulic control system implementing fluidic pressures transmitted
through hydraulic lines to exert or release clamping force between
the two connective surfaces. Operated thusly, the clutch is not
operated in a binary manner, but rather is capable of a range of
engagement states, from fully disengaged, to synchronized but not
engaged, to engaged but with only minimal clamping force, to
engaged with some maximum clamping force. The clamping force
available to be applied to the clutch determines how much reactive
torque the clutch can transmit before the clutch slips.
[0007] The hydraulic control system, as described above, utilizes
lines charged with hydraulic oil to selectively activate clutches
within the transmission. Hydraulic switches or pressure control
solenoids (PCS) are used to selectively apply pressure within a
hydraulic control system. A PCS can be electrically controlled, for
instance with a magnetically actuated solenoid device, well known
in the art. Alternatively, a PCS can be hydraulically controlled,
for example, actuated by a command pressure and a return spring.
Features within the PCS selectively channel or block hydraulic oil
from passing therethrough depending upon the actuation state of the
PCS. In a blocked state, a PCS is known to include an exhaust path,
allowing any trapped hydraulic oil to escape, thereby de-energizing
the connected hydraulic circuit in order to complete the actuation
cycle. Modulation of the command pressure can enable the PCS to be
linearly, variably actuated, including actuation controlling
application of fill pressure to the clutch in order to achieve
within the clutch some middle or transient state between full feed
and exhaust states.
[0008] A hydraulically actuated clutch operates by receiving
pressurized hydraulic oil into a clutch volume chamber. Hydraulic
oil in this clutch volume chamber exerts pressure upon features
within the volume chamber. A piston or similar structure is known
to be utilized to transform this hydraulic pressure into an
articulation, for example a translating motion or compressing
force. In an exemplary hydraulically actuated clutch, pressurized
hydraulic oil is used to fill a clutch volume chamber and thereby
displace a clutch piston in order to selectively apply a
compression force to the connective surfaces of the clutch. A
restoring force, for example as provided by a return spring, is
known to be used to counter the compressive force of the hydraulic
oil. As described above, clutches are known to be engaged through a
range of engagement states. An exemplary clutch with all hydraulic
pressure removed can be in an unlocked state. An exemplary clutch
with maximum hydraulic pressure can be in a locked state. An
exemplary clutch with some partial hydraulic force applied, wherein
the force of the hydraulic oil and the force of a return spring are
substantially equal, the clutch can be in a touching state, with
the plates in contact but with little or no clamping force
applied.
[0009] An engagement of a clutch, accomplished through a clutch
fill event, is known to be accomplished as rapidly as possible,
with some minimum hydraulic pressure being maintained to assure
rapid flow of the hydraulic oil into the clutch volume. However,
rapid engagement of a clutch can cause a perceptible bump in the
vehicle and cause shortened life of the component involved. A shock
absorbing device can be utilized to dampen the force of the rapid
fill of the clutch volume chamber upon the clutch. For example, a
wave plate including a spring feature can be used between the
cylinder piston and the clutch to absorb rapid increases in
hydraulic pressure. The touching state described above can be
defined as the clutch filled with enough hydraulic oil to cause
zero force contact of the wave plate.
[0010] Clutch actuation status, such as would be indicated by a
position sensor on a piston, is frequently not directly monitored.
Such sensors would tend to be expensive, inaccurate, and increase
warranty concerns. However, as described above, clutch operation in
synchronous operation through a plurality of actuation states is a
complicated process including multiple overlapping steps and
control strategies. A method allowing utilization of a command
pressure to precisely actuate a clutch to an important transient
state such as a touching state would be beneficial.
SUMMARY
[0011] A powertrain includes an electro-mechanical transmission
mechanically-operatively coupled to an internal combustion engine
and an electric machine adapted to selectively transmit mechanical
power to an output member via selective application of a plurality
of hydraulically-applied torque transfer clutches. A method to
control the powertrain includes applying through a series of clutch
fill events a series of incrementally changing command pressures in
a pressure control solenoid controllably connected to one of the
clutches, monitoring a pressure switch fluidly connected to the
pressure control solenoid and configured to indicate when the
pressure switch is in a full feed state, determining changes in
cycle times of the pressure switch corresponding to sequential
applications of the series of incrementally changing command
pressures, selecting a preferred command pressure to achieve a
transient state in the clutch based upon the changes in pressure
switch cycle times, and controlling the clutch based upon the
preferred command pressure.
BRIEF DESCRIPTION OF THE DRAWINGS
[0012] One or more embodiments will now be described, by way of
example, with reference to the accompanying drawings, in which:
[0013] FIG. 1 is a schematic diagram of an exemplary powertrain, in
accordance with the present disclosure;
[0014] FIG. 2 is a schematic diagram of an exemplary architecture
for a control system and powertrain, in accordance with the present
disclosure;
[0015] FIG. 3 is a schematic diagram of a hydraulic circuit, in
accordance with the present disclosure;
[0016] FIG. 4 schematically illustrates an exemplary clutch control
circuit utilizing a hydraulically activated pressure control
switch, in accordance with the present disclosure;
[0017] FIG. 5 schematically illustrates an exemplary hydraulically
actuated clutch operated to provide clamping force upon a
mechanical clutch, in accordance with the present disclosure;
[0018] FIG. 6 graphically illustrates fill times required to reach
an overlap state versus the command pressure in the PCS, in
accordance with the present disclosure;
[0019] FIG. 7 graphically illustrates fill times required to reach
an overlap state versus the command pressure through a range of
incrementally decreased command pressures, in accordance with the
present disclosure; and
[0020] FIG. 8 illustrates a flowchart describing an exemplary
process to select and continually validate preferred command
pressures, in accordance with the disclosure.
DETAILED DESCRIPTION
[0021] Referring now to the drawings, wherein the showings are for
the purpose of illustrating certain exemplary embodiments only and
not for the purpose of limiting the same, FIGS. 1 and 2 depict an
exemplary electro-mechanical hybrid powertrain. The exemplary
electro-mechanical hybrid powertrain in accordance with the present
disclosure is depicted in FIG. 1, comprising a two-mode,
compound-split, electro-mechanical hybrid transmission 10
operatively connected to an engine 14 and first and second electric
machines (`MG-A`) 56 and (`MG-B`) 72. The engine 14 and first and
second electric machines 56 and 72 each generate power which can be
transmitted to the transmission 10. The power generated by the
engine 14 and the first and second electric machines 56 and 72 and
transmitted to the transmission 10 is described in terms of input
torques, referred to herein as T.sub.I, T.sub.A, and T.sub.B
respectively, and speed, referred to herein as N.sub.I, N.sub.A,
and N.sub.B, respectively.
[0022] The exemplary engine 14 comprises a multi-cylinder internal
combustion engine selectively operative in several states to
transmit torque to the transmission 10 via an input shaft 12, and
can be either a spark-ignition or a compression-ignition engine.
The engine 14 includes a crankshaft (not shown) operatively coupled
to the input shaft 12 of the transmission 10. A rotational speed
sensor 11 monitors rotational speed of the input shaft 12. Power
output from the engine 14, comprising rotational speed and output
torque, can differ from the input speed, N.sub.I, and the input
torque, T.sub.I, to the transmission 10 due to placement of
torque-consuming components on the input shaft 12 between the
engine 14 and the transmission 10, e.g., a hydraulic pump (not
shown) and/or a torque management device (not shown).
[0023] The exemplary transmission 10 comprises three planetary-gear
sets 24, 26 and 28, and four selectively engageable
torque-transmitting devices, i.e., clutches C1 70, C2 62, C3 73,
and C4 75. As used herein, clutches refer to any type of friction
torque transfer device including single or compound plate clutches
or packs, band clutches, and brakes, for example. A hydraulic
control circuit 42, preferably controlled by a transmission control
module (hereafter `TCM`) 17, is operative to control clutch states.
Clutches C2 62 and C4 75 preferably comprise hydraulically-applied
rotating friction clutches. Clutches C1 70 and C3 73 preferably
comprise hydraulically-controlled stationary devices that can be
selectively grounded to a transmission case 68. Each of the
clutches C1 70, C2 62, C3 73, and C4 75 is preferably hydraulically
applied, selectively receiving pressurized hydraulic oil via the
hydraulic control circuit 42.
[0024] The first and second electric machines 56 and 72 preferably
comprise three-phase AC machines, each including a stator (not
shown) and a rotor (not shown), and respective resolvers 80 and 82.
The motor stator for each machine is grounded to an outer portion
of the transmission case 68, and includes a stator core with coiled
electrical windings extending therefrom. The rotor for the first
electric machine 56 is supported on a hub plate gear that is
operatively attached to shaft 60 via the second planetary gear set
26. The rotor for the second electric machine 72 is fixedly
attached to a sleeve shaft hub 66.
[0025] Each of the resolvers 80 and 82 preferably comprises a
variable reluctance device including a resolver stator (not shown)
and a resolver rotor (not shown). The resolvers 80 and 82 are
appropriately positioned and assembled on respective ones of the
first and second electric machines 56 and 72. Stators of respective
ones of the resolvers 80 and 82 are operatively connected to one of
the stators for the first and second electric machines 56 and 72.
The resolver rotors are operatively connected to the rotor for the
corresponding first and second electric machines 56 and 72. Each of
the resolvers 80 and 82 is signally and operatively connected to a
transmission power inverter control module (hereafter `TPIM`) 19,
and each senses and monitors rotational position of the resolver
rotor relative to the resolver stator, thus monitoring rotational
position of respective ones of first and second electric machines
56 and 72. Additionally, the signals output from the resolvers 80
and 82 are interpreted to provide the rotational speeds for first
and second electric machines 56 and 72, i.e., N.sub.A and N.sub.B,
respectively.
[0026] The transmission 10 includes an output member 64, e.g. a
shaft, which is operably connected to a driveline 90 for a vehicle
(not shown), to provide output power, e.g., to vehicle wheels 93,
one of which is shown in FIG. 1. The output power is characterized
in terms of an output rotational speed, N.sub.O and an output
torque, T.sub.O. A transmission output speed sensor 84 monitors
rotational speed and rotational direction of the output member 64.
Each of the vehicle wheels 93, is preferably equipped with a sensor
94 adapted to monitor wheel speed, V.sub.SS-WHL, the output of
which is monitored by a control module of a distributed control
module system described with respect to FIG. 2, to determine
vehicle speed, and absolute and relative wheel speeds for braking
control, traction control, and vehicle acceleration management.
[0027] The input torques from the engine 14 and the first and
second electric machines 56 and 72 (T.sub.I, T.sub.A, and T.sub.B
respectively) are generated as a result of energy conversion from
fuel or electrical potential stored in an electrical energy storage
device (hereafter `ESD`) 74. The ESD 74 is high voltage DC-coupled
to the TPIM 19 via DC transfer conductors 27. The transfer
conductors 27 include a contactor switch 38. When the contactor
switch 38 is closed, under normal operation, electric current can
flow between the ESD 74 and the TPIM 19. When the contactor switch
38 is opened electric current flow between the ESD 74 and the TPIM
19 is interrupted. The TPIM 19 transmits electrical power to and
from the first electric machine 56 by transfer conductors 29, and
the TPIM 19 similarly transmits electrical power to and from the
second electric machine 72 by transfer conductors 31, in response
to torque commands for the first and second electric machines 56
and 72 to achieve the input torques T.sub.A and T.sub.B. Electrical
current is transmitted to and from the ESD 74 in accordance with
whether the ESD 74 is being charged or discharged.
[0028] The TPIM 19 includes the pair of power inverters (not shown)
and respective motor control modules (not shown) configured to
receive the torque commands and control inverter states therefrom
for providing motor drive or regeneration functionality to meet the
commanded motor torques T.sub.A and T.sub.B. The power inverters
comprise known complementary three-phase power electronics devices,
and each includes a plurality of insulated gate bipolar transistors
(not shown) for converting DC power from the ESD 74 to AC power for
powering respective ones of the first and second electric machines
56 and 72, by switching at high frequencies. The insulated gate
bipolar transistors form a switch mode power supply configured to
receive control commands. There is typically one pair of insulated
gate bipolar transistors for each phase of each of the three-phase
electric machines. States of the insulated gate bipolar transistors
are controlled to provide motor drive mechanical power generation
or electric power regeneration functionality. The three-phase
inverters receive or supply DC electric power via DC transfer
conductors 27 and transform it to or from three-phase AC power,
which is conducted to or from the first and second electric
machines 56 and 72 for operation as motors or generators via
transfer conductors 29 and 31 respectively.
[0029] FIG. 2 is a schematic block diagram of the distributed
control module system. The elements described hereinafter comprise
a subset of an overall vehicle control architecture, and provide
coordinated system control of the exemplary powertrain described in
FIG. 1. The distributed control module system synthesizes pertinent
information and inputs, and executes algorithms to control various
actuators to achieve control objectives, including objectives
related to fuel economy, emissions, performance, drivability, and
protection of hardware, including batteries of ESD 74 and the first
and second electric machines 56 and 72. The distributed control
module system includes an engine control module (hereafter `ECM`)
23, the TCM 17, a battery pack control module (hereafter `BPCM`)
21, and the TPIM 19. A hybrid control module (hereafter `HCP`) 5
provides supervisory control and coordination of the ECM 23, the
TCM 17, the BPCM 21, and the TPIM 19. A user interface (`UI`) 13 is
operatively connected to a plurality of devices through which a
vehicle operator controls or directs operation of the
electro-mechanical hybrid powertrain. The devices include an
accelerator pedal 113 (`AP`) from which an operator torque request
is determined, an operator brake pedal 112 (`BP`), a transmission
gear selector 114 (`PRNDL`), and a vehicle speed cruise control
(not shown). The transmission gear selector 114 may have a discrete
number of operator-selectable positions, including the rotational
direction of the output member 64 to enable one of a forward and a
reverse direction.
[0030] The aforementioned control modules communicate with other
control modules, sensors, and actuators via a local area network
(hereafter `LAN`) bus 6. The LAN bus 6 allows for structured
communication of states of operating parameters and actuator
command signals between the various control modules. The specific
communication protocol utilized is application-specific. The LAN
bus 6 and appropriate protocols provide for robust messaging and
multi-control module interfacing between the aforementioned control
modules, and other control modules providing functionality such as
antilock braking, traction control, and vehicle stability. Multiple
communications buses may be used to improve communications speed
and provide some level of signal redundancy and integrity.
Communication between individual control modules can also be
effected using a direct link, e.g., a serial peripheral interface
(`SPI`) bus (not shown).
[0031] The HCP 5 provides supervisory control of the powertrain,
serving to coordinate operation of the ECM 23, TCM 17, TPIM 19, and
BPCM 21. Based upon various input signals from the user interface
13 and the powertrain, including the ESD 74, the HCP 5 generates
various commands, including: the operator torque request
(`T.sub.O.sub.--.sub.REQ`), a commanded output torque (`T.sub.CMD`)
to the driveline 90, an engine input torque command, clutch torques
for the torque-transfer clutches C1 70, C2 62, C3 73, C4 75 of the
transmission 10; and the torque commands for the first and second
electric machines 56 and 72, respectively. The TCM 17 is
operatively connected to the hydraulic control circuit 42 and
provides various functions including monitoring various pressure
sensing devices (not shown) and generating and communicating
control signals to various solenoids (not shown) thereby
controlling pressure switches and control valves contained within
the hydraulic control circuit 42.
[0032] The ECM 23 is operatively connected to the engine 14, and
functions to acquire data from sensors and control actuators of the
engine 14 over a plurality of discrete lines, shown for simplicity
as an aggregate bi-directional interface cable 35. The ECM 23
receives the engine input torque command from the HCP 5. The ECM 23
determines the actual engine input torque, T.sub.I, provided to the
transmission 10 at that point in time based upon monitored engine
speed and load, which is communicated to the HCP 5. The ECM 23
monitors input from the rotational speed sensor 11 to determine the
engine input speed to the input shaft 12, which translates to the
transmission input speed, N.sub.I. The ECM 23 monitors inputs from
sensors (not shown) to determine states of other engine operating
parameters including, e.g., a manifold pressure, engine coolant
temperature, ambient air temperature, and ambient pressure. The
engine load can be determined, for example, from the manifold
pressure, or alternatively, from monitoring operator input to the
accelerator pedal 113. The ECM 23 generates and communicates
command signals to control engine actuators, including, e.g., fuel
injectors, ignition modules, and throttle control modules, none of
which are shown.
[0033] The TCM 17 is operatively connected to the transmission 10
and monitors inputs from sensors (not shown) to determine states of
transmission operating parameters. The TCM 17 generates and
communicates command signals to control the transmission 10,
including controlling the hydraulic control circuit 42. Inputs from
the TCM 17 to the HCP 5 include estimated clutch torques for each
of the clutches, i.e., C1 70, C2 62, C3 73, and C4 75, and
rotational output speed, N.sub.O, of the output member 64. Other
actuators and sensors may be used to provide additional information
from the TCM 17 to the HCP 5 for control purposes. The TCM 17
monitors inputs from pressure switches (not shown) and selectively
actuates pressure control solenoids (not shown) and shift solenoids
(not shown) of the hydraulic control circuit 42 to selectively
actuate the various clutches C1 70, C2 62, C3 73, and C4 75 to
achieve various transmission operating range states, as described
hereinbelow.
[0034] The BPCM 21 is signally connected to sensors (not shown) to
monitor the ESD 74, including states of electrical current and
voltage parameters, to provide information indicative of parametric
states of the batteries of the ESD 74 to the HCP 5. The parametric
states of the batteries preferably include battery state-of-charge,
battery voltage, battery temperature, and available battery power,
referred to as a range P.sub.BAT_MIN to P.sub.BAT_MAX.
[0035] Each of the control modules ECM 23, TCM 17, TPIM 19 and BPCM
21 is preferably a general-purpose digital computer comprising a
microprocessor or central processing unit, storage mediums
comprising read only memory (`ROM`), random access memory (`RAM`),
electrically programmable read only memory (`EPROM`), a high speed
clock, analog to digital (`A/D`) and digital to analog (`D/A`)
circuitry, and input/output circuitry and devices (`I/O`) and
appropriate signal conditioning and buffer circuitry. Each of the
control modules has a set of control algorithms, comprising
resident program instructions and calibrations stored in one of the
storage mediums and executed to provide the respective functions of
each computer. Information transfer between the control modules is
preferably accomplished using the LAN bus 6 and SPI buses. The
control algorithms are executed during preset loop cycles such that
each algorithm is executed at least once each loop cycle.
Algorithms stored in the non-volatile memory devices are executed
by one of the central processing units to monitor inputs from the
sensing devices and execute control and diagnostic routines to
control operation of the actuators, using preset calibrations. Loop
cycles are executed at regular intervals, for example each 3.125,
6.25, 12.5, 25 and 100 milliseconds during ongoing operation of the
powertrain. Alternatively, algorithms may be executed in response
to the occurrence of an event.
[0036] The exemplary powertrain selectively operates in one of
several operating range states that can be described in terms of an
engine state comprising one of an engine on state (`ON`) and an
engine off state (`OFF`), and a transmission state comprising a
plurality of fixed gears and continuously variable operating modes,
described with reference to Table 1, below.
TABLE-US-00001 TABLE 1 Engine Transmission Operating Applied
Description State Range State Clutches MI_Eng_Off OFF EVT Mode I C1
70 MI_Eng_On ON EVT Mode I C1 70 FG1 ON Fixed Gear Ratio 1 C1 70 C4
75 FG2 ON Fixed Gear Ratio 2 C1 70 C2 62 MII_Eng_Off OFF EVT Mode
II C2 62 MII_Eng_On ON EVT Mode II C2 62 FG3 ON Fixed Gear Ratio 3
C2 62 C4 75 FG4 ON Fixed Gear Ratio 4 C2 62 C3 73
[0037] Each of the transmission operating range states is described
in the table and indicates which of the specific clutches C1 70, C2
62, C3 73, and C4 75 are applied for each of the operating range
states. A first continuously variable mode, i.e., EVT Mode I, or
MI, is selected by applying clutch C1 70 only in order to "ground"
the outer gear member of the third planetary gear set 28. The
engine state can be one of ON (`MI_Eng_On`) or OFF (`MI_Eng_Off`).
A second continuously variable mode, i.e., EVT Mode II, or MII, is
selected by applying clutch C2 62 only to connect the shaft 60 to
the carrier of the third planetary gear set 28. The engine state
can be one of ON (`MII_Eng_On`) or OFF (`MII_Eng_Off`). For
purposes of this description, when the engine state is OFF, the
engine input speed is equal to zero revolutions per minute (`RPM`),
i.e., the engine crankshaft is not rotating. A fixed gear operation
provides a fixed ratio operation of input-to-output speed of the
transmission 10, i.e., N.sub.I/N.sub.O, is achieved. A first fixed
gear operation (`FG1`) is selected by applying clutches C1 70 and
C4 75. A second fixed gear operation (`FG2`) is selected by
applying clutches C1 70 and C2 62. A third fixed gear operation
(`FG3`) is selected by applying clutches C2 62 and C4 75. A fourth
fixed gear operation (`FG4`) is selected by applying clutches C2 62
and C3 73. The fixed ratio operation of input-to-output speed
increases with increased fixed gear operation due to decreased gear
ratios in the planetary gears 24, 26, and 28. The rotational speeds
of the first and second electric machines 56 and 72, N.sub.A and
N.sub.B respectively, are dependent on internal rotation of the
mechanism as defined by the clutching and are proportional to the
input speed measured at the input shaft 12.
[0038] In response to operator input via the accelerator pedal 113
and brake pedal 112 as captured by the user interface 13, the HCP 5
and one or more of the other control modules determine the
commanded output torque, T.sub.CMD, intended to meet the operator
torque request, T.sub.O_REQ, to be executed at the output member 64
and transmitted to the driveline 90. Final vehicle acceleration is
affected by other factors including, e.g., road load, road grade,
and vehicle mass. The operating range state is determined for the
transmission 10 based upon a variety of operating characteristics
of the powertrain. This includes the operator torque request,
communicated through the accelerator pedal 113 and brake pedal 112
to the user interface 13 as previously described. The operating
range state may be predicated on a powertrain torque demand caused
by a command to operate the first and second electric machines 56
and 72 in an electrical energy generating mode or in a torque
generating mode. The operating range state can be determined by an
optimization algorithm or routine which determines optimum system
efficiency based upon operator demand for power, battery state of
charge, and energy efficiencies of the engine 14 and the first and
second electric machines 56 and 72. The control system manages
torque inputs from the engine 14 and the first and second electric
machines 56 and 72 based upon an outcome of the executed
optimization routine, and system efficiencies are optimized
thereby, to manage fuel economy and battery charging. Furthermore,
operation can be determined based upon a fault in a component or
system. The HCP 5 monitors the torque-generative devices, and
determines the power output from the transmission 10 required to
achieve the desired output torque to meet the operator torque
request. As should be apparent from the description above, the ESD
74 and the first and second electric machines 56 and 72 are
electrically-operatively coupled for power flow therebetween.
Furthermore, the engine 14, the first and second electric machines
56 and 72, and the electro-mechanical transmission 10 are
mechanically-operatively coupled to transmit power therebetween to
generate a power flow to the output member 64.
[0039] FIG. 3 depicts a schematic diagram of the hydraulic control
circuit 42 for controlling flow of hydraulic oil in the exemplary
transmission. A main hydraulic pump 88 is driven off the input
shaft 12 from the engine 14, and an auxiliary pump 110 controlled
by the TPIM 19 to provide pressurized fluid to the hydraulic
control circuit 42 through valve 140. The auxiliary pump 110
preferably comprises an electrically-powered pump of an appropriate
size and capacity to provide sufficient flow of pressurized
hydraulic oil into the hydraulic control circuit 42 when
operational. The hydraulic control circuit 42 selectively
distributes hydraulic pressure to a plurality of devices, including
the torque-transfer clutches C1 70, C2 62, C3 73, and C4 75, active
cooling circuits for the first and second electric machines 56 and
72 (not shown), and a base cooling circuit for cooling and
lubricating the transmission 10 via passages 142, 144 (not depicted
in detail). As previously stated, the TCM 17 actuates the various
clutches to achieve one of the transmission operating range states
through selective actuation of hydraulic circuit flow control
devices comprising variable pressure control solenoids (`PCS`) PCS1
108, PCS2 114, PCS3 112, PCS4 116 and solenoid-controlled flow
management valves, X-valve 119 and Y-valve 121. The hydraulic
control circuit 42 is fluidly connected to pressure switches PS1,
PS2, PS3, and PS4 via passages 122, 124, 126, and 128,
respectively. The pressure control solenoid PCS1 108 has a control
position of normally high and is operative to modulate the
magnitude of fluidic pressure in the hydraulic circuit through
fluidic interaction with controllable pressure regulator 107 and
spool valve 109. The controllable pressure regulator 107 and spool
valve 109 interact with PCS1 108 to control hydraulic pressure in
the hydraulic control circuit 42 over a range of pressures and may
provide additional functionality for the hydraulic control circuit
42. Pressure control solenoid PCS3 112 has a control position of
normally high, and is fluidly connected to spool valve 113 and
operative to effect flow therethrough when actuated. Spool valve
113 is fluidly connected to pressure switch PS3 via passage 126.
Pressure control solenoid PCS2 114 has a control position of
normally high, and is fluidly connected to spool valve 115 and
operative to effect flow therethrough when actuated. Spool valve
115 is fluidly connected to pressure switch PS2 via passage 124.
Pressure control solenoid PCS4 116 has a control position of
normally low, and is fluidly connected to spool valve 117 and
operative to effect flow therethrough when actuated. Spool valve
117 is fluidly connected to pressure switch PS4 via passage
128.
[0040] The X-Valve 119 and Y-Valve 121 each comprise flow
management valves controlled by solenoids 118, 120, respectively,
in the exemplary system, and have control states of High (`1`) and
Low (`0`). The control states refer to positions of each valve to
which control flow to different devices in the hydraulic control
circuit 42 and the transmission 10. The X-valve 119 is operative to
direct pressurized fluid to clutches C3 73 and C4 75 and cooling
systems for stators of the first and second electric machines 56
and 72 via fluidic passages 136, 138, 144, 142 respectively,
depending upon the source of the fluidic input, as is described
hereinafter. The Y-valve 121 is operative to direct pressurized
fluid to clutches C1 70 and C2 62 via fluidic passages 132 and 134
respectively, depending upon the source of the fluidic input, as is
described hereinafter. The Y-valve 121 is fluidly connected to
pressure switch PS1 via passage 122.
[0041] The hydraulic control circuit 42 includes a base cooling
circuit for providing hydraulic oil to cool the stators of the
first and second electric machines 56 and 72. The base cooling
circuit includes fluid conduits from the valve 140 flowing directly
to a flow restrictor which leads to fluidic passage 144 leading to
the base cooling circuit for the stator of the first electric
machine 56, and to a flow restrictor which leads to fluidic passage
142 leading to the base cooling circuit for the stator of the
second electric machine 72. Active cooling of stators for the first
and second electric machines 56 and 72 is effected by selective
actuation of pressure control solenoids PCS2 114, PCS3 112 and PCS4
116 and solenoid-controlled flow management valves X-valve 119 and
Y-valve 121, which leads to flow of hydraulic oil around the
selected stator and permits heat to be transferred therebetween,
primarily through conduction.
[0042] An exemplary logic table to accomplish control of the
exemplary hydraulic control circuit 42 to control operation of the
transmission 10 in one of the transmission operating range states
is provided with reference to Table 2, below.
TABLE-US-00002 TABLE 2 X- Y- Transmission Valve Valve Operating
Logic Logic PCS1 PCS2 PCS3 PCS4 Range No C2 Normal Normal Normal
Normal State Latch Latch High High High Low EVT 0 0 Line MG-B C1
MG-A Mode I Modulation Stator Stator Cool Cool EVT 0 1 Line C2 MG-B
MG-A Mode II Modulation Stator Stator Cool Cool Low 1 0 Line C2 C1
C4 Range Modulation High 1 1 Line C2 C3 C4 Range Modulation
[0043] A Low Range is defined as a transmission operating range
state comprising one of the first continuously variable mode and
the first and second fixed gear operations. A High Range is defined
as a transmission operating range state comprising one of the
second continuously variable mode and the third and fourth fixed
gear operations. Selective control of the X-valve 119 and the
Y-valve 121 and actuation of the solenoids PCS2 112, PCS3 114, PCS4
116 facilitate flow of hydraulic oil to actuate clutches C1 70, C2
63, C3 73, and C4 75, and provide cooling for the stators the first
and second electric machines 56 and 72.
[0044] In operation, a transmission operating range state, i.e. one
of the fixed gear and continuously variable mode operations, is
selected for the exemplary transmission 10 based upon a variety of
operating characteristics of the powertrain. This includes the
operator torque request, typically communicated through inputs to
the UI 13 as previously described. Additionally, a demand for
output torque is predicated on external conditions, including,
e.g., road grade, road surface conditions, or wind load. The
operating range state may be predicated on a powertrain torque
demand caused by a control module command to operate of the
electrical machines in an electrical energy generating mode or in a
torque generating mode. The operating range state can be determined
by an optimization algorithm or routine operable to determine an
optimum system efficiency based upon the operator torque request,
battery state of charge, and energy efficiencies of the engine 14
and the first and second electric machines 56 and 72. The control
system manages the input torques from the engine 14 and the first
and second electric machines 56 and 72 based upon an outcome of the
executed optimization routine, and system optimization occurs to
improve fuel economy and manage battery charging. Furthermore, the
operation can be determined based upon a fault in a component or
system.
[0045] FIG. 4 schematically illustrates an exemplary clutch control
circuit utilizing a hydraulically activated pressure control
switch, in accordance with the present disclosure. Clutch control
circuit 200 includes PCS 210, pressure switch 240, and hydraulic
lines 270, 272, 274, 276, 278, and 280. PCS 210 selectively
controls flow of pressurized hydraulic oil to and from a
hydraulically actuated clutch (not shown) by translation of
selecting mechanism within the PCS, in this exemplary embodiment, a
spool valve plunger 220. Plunger 220 is selectively acted upon from
a first end 222 of the plunger and a second end 224 of the plunger,
the balance of forces determining the translative position of the
plunger within the PCS. Plunger 220 includes plunger details 226
including holes, grooves, channels, or other features formed on the
plunger in order to selectively direct hydraulic oil between
various ports connecting hydraulic lines to PCS 210. The position
of plunger 220 within PCS 210, corresponding to clutch states
described above, selectively align plunger details 226 with
hydraulic lines accomplishing the intended clutch function. In the
exemplary clutch of FIG. 4, a plunger position to the right
corresponds to a full feed state, wherein hydraulic pressure from a
main pressure line 272 is channeled through plunger details 226 to
clutch feed line 276. Similarly, a plunger position to the left
corresponds to an exhaust state, wherein hydraulic oil within the
clutch is allowed to escape the clutch and flow through exhaust
line 274, entering a hydraulic control system return line (not
shown). Selecting the position of plunger 220 is accomplished by
modulating a command pressure to a command pressure line 270
feeding a command pressure volume 260 in contact with first end
222. As will be appreciated by one having ordinary skill in the
art, force created by pressure on a surface can be determined
through the following equation:
FORCE=PRESSURE*SURFACE_AREA_ACTED_UPON [1]
In the case of exemplary plunger 220, the force acting upon the
plunger from the left equals the hydraulic pressure achieved within
command pressure volume 260 times the surface area of first end
222. An increase in pressure within command pressure volume 260
increases the force acting upon plunger 220 from the side of first
end 222. A valve return spring 250 applies a force to the second
end 224, acting as a restorative force in the opposite direction of
the pressure within command pressure volume 260. Force resulting
from pressure within volume 260 and force from spring 250 act
together such that increased pressure within command pressure
volume 260 tends to move plunger 220 in one direction, and reduced
pressure within command pressure volume 260 tends to move plunger
220 in the opposite direction. Exemplary PCS 210 includes another
feature including a feedback line 278. Hydraulic oil flowing
through clutch feed line 276 additionally flows or applies a
pressure through feedback line 278. Hydraulic oil from feedback
line 278 re-enters PCS 210 within a feedback pressure volume 265
located on the same side of plunger 220 as spring 250. Force
resulting upon plunger 220 from hydraulic pressure within feedback
pressure volume 265 counteracts force resulting from hydraulic
pressure within command pressure volume 260. As a result, wherein a
balance of forces resulting from pressure within command pressure
volume 260 and spring 250 would cause plunger 220 to be in a
position correlating to a full feed state, elevated pressure
achieved within clutch feed line 276 associated with a clutch fill
event reaching a certain progression creates a force acting upon
plunger 220 away from the full feed state position. Calibration
and/or control of feedback line 278 and resulting force upon
plunger 220 corresponding to a selected pressure within command
pressure volume 260 can create a self-correcting plunger position
between the opposite ends of plunger travel, enabling an overlap
state. Such an overlap state is useful for modulating the pressure
achieved within the clutch, for example, enabling calibrated
control to a touching state for the clutch. Full feed state can
still be achieved despite operation of the feedback line 278 by
setting pressure within the command pressure volume 260 to apply a
force to plunger 220 exceeding the combination of the force applied
by spring 250 and force resulting from hydraulic pressure within
feedback pressure volume 265. PCS 210 is known to include pressure
switch 240, fed by pressure switch line 280, utilized in known
control methods to indicate pressure levels required for control of
PCS 210. In this way, PCS 210 can selectively channel hydraulic oil
to accomplish multiple states within a hydraulically activated
clutch.
[0046] By modulating a command pressure, a PCS of the above
exemplary configuration can operate in three states. A high command
pressure commands a full feed state, allowing full exposure of
P.sub.LINE to the clutch being filled. A low or null command
pressure commands an exhaust state, blocking access of P.sub.LINE
to the clutch and providing a path to exhaust hydraulic pressure
from within the clutch. An intermediate or calibrated command
pressure commands an overlap state. The function of an overlap
state depends upon the calibration of the calibrated command
pressure. An exemplary function of such an overlap state is to
command a touching state in the clutch. Selective calibration of
the command pressure to achieve the overlap state, in combination
with monitored operation of the pressure switch, allows for
accurately selecting a fill level within the clutch, for example, a
fill level corresponding to a touching state in the clutch.
[0047] A number of PCS physical configurations are known. One
exemplary PCS configuration, as described above, utilizes a
cylindrical plunger located in a cylindrical housing. However, a
multitude of shapes, configurations, activations methods, and
calibration strategies are known in the art, and this disclosure is
not intended to be limited to the particular exemplary embodiments
described herein.
[0048] Pressure switch 240 is calibrated to indicate a pressure
reaching some level. In the particular embodiment described in FIG.
4, the pressure switch can be utilized for example, to indicate a
positive signal only when the PCS is in a full feed state. In such
an exemplary use, the calibration of the pressure switch indication
need not correspond to the actual pressures to which it is exposed,
for example pressure levels in the command pressure volume 260, but
can rather indicates some nominal level which the pressure always
exceeds when the pressure switch is exposed to the pressurized
hydraulic fluid.
[0049] As described above, operation and control of clutches are
important to operating a complex powertrain, such as a hybrid
powertrain. Drivability, fuel efficiency, and component life are
all impacted by the operation of clutches within the system. Known
methods utilizing look-up tables to control clutch activating
devices, such as a PCS, are imprecise and inefficient. Much can be
determined within a hydraulic control system based upon analysis of
available inputs. A method is disclosed for localizing a preferred
command pressure to attain a touching state within a transmission
clutch based upon clutch fill times and pressure switch
readings.
[0050] A hydraulically actuated clutch utilizes selectively
actuated pressurized hydraulic flow to create a desired motion or
compression. An exemplary clutch operates by receiving pressurized
hydraulic oil into a clutch volume chamber. FIG. 5 schematically
illustrates an exemplary hydraulically actuated clutch operated to
provide clamping force upon a mechanical clutch, in accordance with
the present disclosure. Clutch assembly 300 comprises a clutch
cylinder 320 and a mechanical clutch 340. Clutch cylinder 320
includes a piston 322 and a clutch volume chamber 324. Pressurized
hydraulic fluid at some fill pressure enters clutch volume chamber
324 through hydraulic line 350. Hydraulic line 350 is fluidly
connected with a mechanism for selectively applying hydraulic flow,
such as an exemplary PCS device (not shown). Hydraulic oil in
clutch volume chamber 324 exerts pressure upon features within the
volume chamber. Piston 322 transforms the fill pressure exerted by
the hydraulic fluid into a force. The force transmitted through
piston 322 is used to articulate mechanical clutch 340 through
various states required according to synchronous clutch operation
described above. Positive hydraulic pressure is used to fill the
clutch volume chamber 324 and move piston 322 in one direction. As
will be appreciated by one having ordinary skill in the art,
evacuation of hydraulic oil from clutch volume chamber 324 acts in
some degree to move piston 322 in the other direction, but
cavitation limits the ability of low pressure hydraulic fluid from
effectively moving piston 322. As a result, return spring 326 is
utilized to provide force to move piston 322 in the direction
opposite to the direction achieved through the application of
pressurized hydraulic fluid.
[0051] Mechanical clutch 340 is selectively actuated by the
transmission of force through piston 322. Mechanical clutch 340
includes clutch connective surfaces in the form of clutch plates
345. Clutch plates 345 are connected to rotating members within the
transmission (not shown). When mechanical clutch 340 is not
actuated, clutch plates 345 are kept separate. Spinning of some
fraction of clutch plates 345 does not cause spinning of the
remaining fraction of clutch plates 345. When mechanical clutch 340
is actuated, clutch plates 345 are brought into contact with
neighboring plates, and sufficient frictional forces between clutch
plates 345 creates a locked relationship wherein the plates move in
unison. Between rotating objects applying a torque, the torque
capacity (`T.sub.C`) generated between the objects can be
determined by the following equation:
T C = 2 3 * f * F A [ 2 ] ##EQU00001##
f is the coefficient of friction between the rotating objects. As
will be appreciated by one having ordinary skill in the art, f
changes depending upon whether there is relative movement between
the two objects. F.sub.A is the axial force applied normally to
direction of rotation of the objects. F.sub.A in mechanical clutch
340 is generated by compressive force transmitted through piston
322. When the clutch is in a touching state, F.sub.A is kept at
substantially zero, yielding zero torque capacity.
[0052] A process transitioning piston 322 from one extreme of
motion to the other includes three distinct phases. A first phase,
beginning from a fully unlocked state in the clutch, wherein no
hydraulic pressure is being applied upon piston 322, the exemplary
piston 322 is in a fully left position, as depicted in FIG. 5, and
has no contact with mechanical clutch 340 or clutch plates 345. As
pressurized hydraulic fluid at a fill pressure is directed into
clutch volume chamber 324, force is applied to the piston and it
begins to move to the right. Because piston 322 is not yet in
contact with mechanical clutch 340, piston 322 moves relatively
easily, and pressurized fluid entering clutch volume chamber 324
achieves a relatively rapid movement of piston 322. During this
first phase, the volume of hydraulic fluid in the clutch volume
chamber 324 changes rapidly. A third phase can be defined once
piston 322 moves toward the right extreme of travel and contacts
mechanical clutch 340, force applied upon piston 322 transmits
force to mechanical clutch 340 and creates compressive pressure
between clutch plate 345. Because piston 322 is subject to equal
force from clutch plates 345 as the piston is transmitting, piston
322 moves much more slowly as a result of pressurized fluid acting
upon the piston. In this third phase, because the piston only moves
with additional compression of components of mechanical clutch 340,
the volume of hydraulic fluid in clutch volume chamber 324 changes
more slowly. In the transitional period between the piston in zero
contact and the application of compressive force upon the clutch
plates, a second phase can be defined wherein piston 322
transitions from a period of relatively rapid movement and a period
of relatively slower movement. Abrupt application of force upon
clutch plates 345 can have adverse effects, including damage to the
plates and potential perception of the contact. Wave plate 328 can
be used as part of mechanical clutch 340 to absorb some portion of
force of the abrupt contact, making the transition between the
first and third phases less abrupt. Further, a touching state can
be defined wherein between the end of the first phase and the
initiation of the second phase, wherein the piston begins to
contact and exert force upon mechanical clutch 340 and wave plate
328.
[0053] As described above, clutches transition between locked and
unlocked states, and clutches designed to operate synchronously or
without slip require substantially zero relative velocity when
reactive torque is transmitted through the clutch. Strategies for
synchronous operation of clutches include synchronizing the clutch
connective surfaces, then applying a clamping force to lock the
clutch, thereby creating a clutch torque capacity in the clutch,
and then transmitting reactive torque through the clutch.
[0054] Clutch control strategies, sequentially, and in some
instance simultaneously, performing operations to synchronize the
clutch plates, actuate the clutch to first the touching state and
then to a fully locked state, and then ramp up torque capacity of
the locked clutch. The order in which these operations are
performed are important to synchronous operation, but also, the
entire clutch transition must occur in as short a time span as
possible to preserve drivability. Commands must be given to various
powertrain components, accounting for reaction times, in order to
generate the various operations involved in a shift occur in order
with as little delay as possible. Commands and resulting actions
can be simultaneous and overlapping, and understanding the time
that various components take to reach a particular state in
response to commands is important to coordinating the reactions in
the order required in synchronous clutch operation. Commands to the
hydraulic control system actuating the clutch and the resulting
actions in the clutch are important to the sequential steps
described above.
[0055] A pressure switch cycle time is a measure of the response
times that result in clutch fill operations from the initiation of
a hydraulic flow to a hydraulically actuated clutch until some
clutch state of interest, as established by the configuration of
the PCS controlling clutch operation. Pressure switch cycle times,
measuring a time to a PCS entering a full feed state to exiting the
full feed state, wherein an overlap state following the full feed
state is intended to create a touching state in an associated
clutch, can be utilized to diagnose a time until the clutch reaches
a touching state through analysis of the times. An exemplary method
to utilize pressure switch cycle times is to track, first, a
pressure switch signal, corresponding to a command pressure
initiating a full feed state and exceeding the calibrated pressure
of the pressure switch, and, second, a pressure switch signal,
corresponding to a drop in sensed pressure, for example, if the
command pressure is cut off from the pressure sensor through the
plunger of the PCS reaching an overlap state. The time span between
these two pressure switch signals can be tracked as a pressure
switch cycle time measuring the time required to create an overlap
state in the PCS. FIG. 6 graphically illustrates fill times
required to reach an overlap state versus command pressure in a
PCS, in accordance with the present disclosure.
[0056] The exemplary data of FIG. 6 demonstrates an overall trend
in the pressure switch cycle times. Fill pressures resulting in
clutch feed line and within a clutch volume chamber are the result
of P.sub.LINE applied to the PCS minus any pressure losses
resulting from flow through the PCS and relates lines. Pressure
losses resulting from flow can be described by the following
equation:
PRESSURE_LOSS=FLOW*FLOW_RESISTANCE [3]
Flow resistance is a fixed term for a fixed geometry of the PCS at
a given setting. Pressure loss is therefore proportional to flow.
Flow through the PCS to the clutch is high when the clutch piston
is being displaced, for example, in the first phase described
above. Flow through the PCS to the clutch is low when the clutch
piston is relatively stationary, for example, in the third phase
described above, wherein the piston is actively compressing the
clutch plates. Applied to the trend in the data of FIG. 6, low
command pressures correspond to low pressure switch cycle times.
Pressure switch cycle times relate the time between the initiation
and end of the full feed state. A low command pressure is easily
countered by the feedback pressure resulting in the PCS from the
application of P.sub.LINE in the full feed state. The resulting
fill pressure in the clutch has only started to ramp up when the
PCS is set to the overlap state, arresting any further increase in
pressure within the clutch cylinder, so the PCS reaches the overlap
state while the clutch is still in the first phase, described
above. As the command pressures increase, the pressure switch cycle
times also quickly increase. Because the clutch is still in the
first phase described above wherein the piston is quickly displaced
by hydraulic flow, the resulting high flow still results in high
pressure losses. As a result, the fill pressure and the resulting
pressure within the feedback loop rise slowly, resulting in
significantly long incremental pressure switch cycle times with
each incremental increase in the command pressure. Once the piston
has been displaced and the clutch enters the third phase, described
above, additional force applied to the piston through application
of fill pressure produces less movement, corresponding to
compression of the wave plate and clutch plates. Volume in the
clutch volume chamber changes slowly, and the resulting hydraulic
flow into the cylinder is reduced. As a result, pressure losses
resulting from flow decrease, and the fill pressure rapidly
approaches a static pressure or P.sub.LINE. Because the fill
pressure in this third phase increases rapidly, the time span
needed to incrementally increase fill pressure decreases. Increases
in command pressure at higher command pressures result in only
small increases in pressure switch cycle times.
[0057] As a result of the above behavior, low command pressures to
the PCS correspond to low pressure switch cycle times. As command
pressures increase, the incremental times to reach an overlap state
in the PCS increase quickly at first, and then more slowly as the
clutch plates begin to compress. The transition between these two
regions of pressure switch cycle time behavior describes the
transition between the first phase and third phase, or the second
phase. As described above, the touching state occurs between the
end of the first phase and the initiation of the second phase.
[0058] As described above, command pressures in the steep section
of the curve correspond to a PCS ceasing full feed state operation
before the touching state in the connected clutch is achieved. In a
clutch shift requiring that a touching state be efficiently
achieved, transitioning the PCS to an overlap state before the
touching state is reached is not preferable. However, high fill
pressures overshoot the touching state in the clutch and can cause
drivability issues. By analyzing a sample of pressure switch cycle
times through a range of command pressures, differences in the
cycle times for incremental increases in the command pressure can
be used to calibrate or determine a preferred command pressure to
quickly and precisely produce a touching state.
[0059] FIG. 7 graphically illustrates fill times required to reach
an overlap state versus command pressure in a PCS through a range
of incrementally decreased command pressures, in accordance with
the present disclosure. As described above in relation to FIG. 6,
command pressures in the steep section of the curve correspond to a
PCS ceasing full feed state operation before the touching state in
the connected clutch is achieved. Because this result in operation
is not preferable and because overly high command pressures can be
anticipated and adjusted for, a method to detect a transition in
incremental pressure switch cycle times beginning with slightly
high command pressures is preferred. Command pressures are shown as
incrementally decreasing samples P.sub.1 through P.sub.6.
Corresponding pressure switch cycle times are shown as T.sub.1
through T.sub.6. Changes between the pressure switch cycle times
are depicted as D.sub.1 through D.sub.5. A comparison of D values
yields little change between D.sub.1 and D.sub.3, with these values
belonging to a subset of D values with the shortest changes to
pressure switch cycle times. However, D.sub.4 is significantly
increased from D.sub.3. This change indicates that a transition in
pressure switch cycle times is indicated between P.sub.4 and
P.sub.5. By selecting a preferred command pressure at or above
P.sub.4, the command pressure will not cease a full feed state
prior to the touching state being achieved. One preferred method,
to insure that the preferred command pressure is robustly above the
transition point indicated in FIG. 7, is to add a command pressure
adjustment to the first P value showing the increased D value. An
exemplary command pressure increase is one and a half times the
value of the incremental decrease utilized in the calibration
samples. In the present exemplary data, this would create a
preferred command pressure half way between P.sub.4 and P.sub.3. In
this way, pressure switch cycle times can be compared through a
range of command pressures to calibrate or determine a preferred
command pressure to efficiently achieve a touching state.
[0060] The above methods to calibrate a preferred command pressure
can be performed once and maintained indefinitely for use in
filling the clutch. However, with changing temperatures and wear in
the system over time, behavior of the clutch through a fill event
can change. Characteristics of clutch fill events, such as measured
pressure switch cycle times versus expected pressure switch cycle
times, can be used to continually or periodically validate the
preferred command pressure. In the event that the measure values
differ from expected values by more than a threshold, the
calibration process can be reinitiated to determine a new preferred
command pressure. This process can occur a number of times through
the lifespan of a powertrain in order to maintain an ability to
precisely indicate a touching state in a clutch fill event.
[0061] FIG. 8 illustrates a flowchart describing an exemplary
process to select and continually validate preferred command
pressures, in accordance with the disclosure. Process 400 starts in
step 402. At step 404, an untrained state is initiated. At step
406, an old preferred command pressure is purged, if it exists from
a previous calibration. At step 408, a training state is initiated.
At step 410, an iteratively decreasing command pressure calibration
is performed, according to methods described herein. At step 412, a
preferred command pressure indicating occurrence of a touching
state is selected, according to methods described herein. At step
414, a trained state is initiated. At step 416, measured pressure
switch cycle times are compared to expected pressure switch cycle
times. At step 418, the comparison of 416 is utilized to either
validate or invalidate the preferred command pressure. If the
preferred command pressure remains validated, then the process
reiterates to 416 wherein future measured fill times are compared
to expected fill times. If the preferred command pressure is
invalidated, then the process reiterates to step 404, wherein the
selection of a new preferred command pressure begins.
[0062] The methods described herein can be performed in a PCS
control module located within a larger control system or located
individually as a unitary device.
[0063] It is understood that modifications are allowable within the
scope of the disclosure. The disclosure has been described with
specific reference to the preferred embodiments and modifications
thereto. Further modifications and alterations may occur to others
upon reading and understanding the specification. It is intended to
include all such modifications and alterations insofar as they come
within the scope of the disclosure.
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